Impact tools



F. A. KAMAN 2,753,965

IMPACT TooLs 2 Sheets-Sheet l July 10, 1956 Filed oct. s, 1951 ///y mb mmm m. @QM mm NM@ ATM A@ Q@ ww um uw .m |T Lkw QN JUIY 10 1956 F. A. KAMAN 2,753,965

IMPACT TOOLS Filed Oct. 3, 1951 2 Sheets-Sheet 2 im@ yy INVENToR. fran/6 /.fzmmf United States Patent O IMPACT TOOLS Frank A. Kaman, Aurora, Ill., assignor to Thor Power Tool Company, a corporation of Delaware Application October 3, 1951, Serial No. 249,546

6 Claims. (Cl. 1922-305) This invention relates to improvements in portable power operated impact tools for setting and removing nuts and bolts by power, and more particularly to improvements in the cam actuated driving mechanism employed in such impact tools for clutching and declutching the hammer and anvil members and for imparting rotational hammer blows on the anvil member to tighten or loosen a nut or bolt.

In the copending application Serial No. 81,966, tiled March 17, 1949, by Walter G. Mitchell, now U. S. Patent No. 2,585,486, issued February 12, 1952, there is described and claimed an impact clutch mechanism including a novel arrangement of a pair of rotatable hammer and anvil members, impact means, and slidable cam means for actuating the impact means. The present invention involves a novel and improved form of cam means utilizing a pair of rotatable cam elements operatively interconnected by a plurality of rolling elements such as balls. As will appear hereinafter, the invention is particularly characterized by an improved arrangement of the cam elements and the coacting balls whereby the balls are effectively supported at all times and are not subject to undesirable shearing stress. It will be readily understood that in an impact tool it is highly important to avoid or minimize shear conditions because of the danger of failure of the parts in shear under the repeated impact forces.

Accordingly, the primary object of my invention is to provide an impact clutch mechanism having novel and improved cam means for actuating the impact clutch mechanism.

A further object of the invention is to provide in an impact clutch mechanism a novel actuating means including a plurality of ball cam elements in which the ball elements are not subjected to excessive shearing stress during operation of the impact clutch mechanism.

Other and further objects and advantages of the present invention will become apparent hereinafter as the description progresses, reference being had to the accompanying drawings in which:

Fig. 1 is a longitudinal sectional view with portions thereof` in elevation and showing an impact tool comprising a preferred embodiment of the present invention;

Fig. 2 is a fragmentary longitudinal sectional view of the impact tool taken on the line 2 2 in Fig. 1 and showing the device in one of its operating positions;

Fig. 3 is a fragmentary sectional view showing portions of the device in a different operating position;

Fig. 4 is a transverse sectional View taken on the line 4 4 in Fig. 1;

Fig. 5 is a transverse sectional view as taken on the line 5 5 in Fig. 1 with portions of the structure shown in elevation;

Fig. 6 is a transverse sectional view taken on the line Fig. 8 is a fragmentary sectional View taken on the line 8 8 of Fig. 3; and

Fig. 9 is a perspective exploded View of a portion of" the tool in disassembled position.

Referring now to the drawings, the impact tool illustrated therein comprises a four section housing including an impact unit section 20 at the forward portion of the tool, a gear section 20a at the intermediate portion thereof, a field case 21 containing an electric motor, not shown, and a cap 22 at the rear end thereof. As more fully described in the earlier application Serial No. 81,966, the field case 21 is provided at its rearward end with supporting structure for a ball bearing unit having journaled therein the rear end of an armature shaft 26. A motor reversing ring 28, which may be composed of plastic or other suitable material, abuts the extreme rearward end of the housing section 21 and is held in position thereagainst for limited rotative movement by the cap 22. As explained more fully in the earlier application, relative rotative movement of the plastic ring 28 in one direction or the other serves to reverse the operational direction of the motor contained in the housing section 21. The armature shaft 26 is journaled on its forward side in a ball bearing unit 37 supported within a bushing 38 which is secured within the axial opening of an annular, cup-shaped cross portion 39 extending diagonally inwardly and rearwardly from the forward end of the housing section 20a.

Rotary motion of the electric motor contained in the housing section 21 is transmitted to the impact unit within the forward housing section 2t) by means of a planetary gear train. The forward end of the armature shaft 26 is shaped to provide a gear 46 in constant mesh with a pair of diametrically opposed, planetary gears 47 mounted for free rotative movement on short shafts 48 secured ina spider member 49. The spider member 49 is integrally formed as the upper or rearward end portion of a drive shaft 50 and its extreme rearward end is journaled for` free rotative movement in a thrust ball bearing unit 51 also received in the bushing 38. The planetary gears 47 are also in constant mesh with an internal ring gear 52 secured in and projecting somewhat beyond the extreme forward open end of the housing section 20a. The ring gear 52 is provided with an outwardly directed annular shoulder 53 which forms a seat against which the rearward end of the impact housing section 20 may be drawn up into tight engagement by means of suitable screws,

gear 52 out of meshing engagement therewith. As will Y become apparent hereinafter, the impact mechanism may then be removed rearwardly out of the housing section 2t).

The forward end of the drive shaft 50, comprising the main supporting member or stem of the impact unit, is

received within an axial bore 54 of an impact spindle or anvil 56 which terminates in a square section 57 exteriorly of the housing section 2.0 and is adapted to mount and drive a nut or bolt engaging socket (not shown) in the usual manner. The impact spindle 56 is adapted to rotate within a bushing 58 secured by a set screw 59 within the axial opening at the forward end of the housing section 20. A rotatable hammer member 60 is mounted on the drive shaft 50 above the impact spindle 56 for free and independent rotative movement relative thereto, the hammer 60 being provided with an axially extending central opening 55' in which is secured a needle bearing 61 facilitating rotative movement between the parts. A thrust washer 62 is also mounted on the shaft 50 between the hammer 60 and the impact spindle 56.

The hammer 60 is provided with longitudinally extending openings 63 on diametrically opposite sides of the central axial opening 55, the openings 55 and 63 thus being parallel and in alignment, and the hammer 6i) thus being of solid, rugged construction and possessing great mass and inherently great strength. A pair of impacting rods or jaws 64 are slidably mounted within and extend through the hammer openings 63, the lower ends of the rods 64 being adapted to project when in impacting position, as shown in Fig. 1, beyond the forward end surface of the hammer 60 and into arcuate slotted openings 66 in a radially outwardly directed iianged portion 67 of the anvil 56. As will be explained in detail hereinafter, the impact rods 64 rotate with the hammer 60` and are adapted to strike the anvil 56 at the abutments formed by the ends of the arcuate slots 66 to cause rotation of the anvil.

The upper ends of the impact rods 64 are each provided with an annular groove 68 into which are received the edges of bifurcated portions 69 extending radially from the forward end of an arched cam lifter member or plate 70. The impact rods 64 may be readily mounted on the cam lifter plate 70 by sliding them inwardly into the slots 65 thereof from the outer edges of the bifurcated portions 69 and they are then free to rotate relative thereto. The cam lifter plate '70 is provided with a central axial opening through which the drive shaft 50 extends, the plate 70 being slidable on the shaft 50 for axial movement relative thereto. A relatively heavy compression spring 71 encircles the cam plate 70 and the forward end thereof seats against the radial edge portions of the bifurcated portions 69 outwardly of the heads of the impact rods 64. By this arrangement the cam plate 70 is constantly urged forwardly so as to project the impact rods 60 into impacting position. The spring 71 also serves the further purpose of retaining the impact rods 60 in assembled position within the slots 65 of the bifurcated portions 69 during assembly of the various parts of the tool and permits the impact mechanism to be assembled and handled as an integral unit,

The upper end of the compression spring 71 bears and seats against radially directed flanges 72 of an upper spring plate 73, this plate 73 also being slidable on the drive shaft 50 and being separated from the forward end surface of the spider member 49 by thrust bearings 74. 'Ihe upper spring plate 73 is formed with a hub portion around the shaft 50 on the forward side thereof to provide an end surface 76 for engagement with a flat cooperating surface 77 on the rearward end of the cam lifter plate 70, the surface 77 being adapted to be moved upwardly into such engagement upon extreme rearward movement of the lift plate 70.

As will hereinafter appear in connection with the operation of the tool, the cam lifter plate 70 is shiftable axially of the shaft 50 and against the force of the spring 71 1n response to a predetermined resistance torque tending to prevent rotation of the anvil 56. The rotary driving connection between the cam lifter plate 70 and the drive shaft 50 and also the actuating means for shifting the cam lifter plate 70 axially on the shaft S0 are both provided by means of a cam mechanism consisting of a pair of rotatable driving and driven cam elements and a plurality of coacting ball elements. For this purpose, the cam lifter plate 70 is formed with a generally cup-shaped central portion having suitable cam surfaces at Vits forward or under side and constituting the driven member of the aforementioned pair of rotatable cam elements. A coacting drive cam member or bushing 78 is secured to the shaft 50 adjacent the forward side of the cam lifter plate 70 and is formed with complementary-cam surfaces operatively engaging the cam surfaces of the member 70 through a pair of ball elements 79.

Referring particularly to Figs. l and 5, it will be seen that the interior cam surfaces of the member 70 are provided at the inner face of the base of the cup-shaped member and comprise a pair of oppositely disposed generally V-shaped cam tracks or grooves 80 which extend in arcuate paths along the curved side walls or skirt portion,

indicated at 81, of the central cup-shaped portion of the member 70. Extending between the adjacent ends of the tracks or grooves 80 are a pair of flat raised circumferential bearing surfaces 82 merging with the grooves 80, and

`The lowermost points, indicated at 84 (Fig. 5), of each of the grooves or channels 80 are adapted to receive the ball elements 79 in the normal driving position of the cam means.

The drive cam member or bushing 78 is keyed to the shaft S0 and restrained against axial movement thereon by means of a pair of retaining balls 86 (Figs. 2 and 4) which are received in suitable aligned cavities in the member 78 and the shaft 50. The forward end of the drive cam member 78 is received within an annular recess 87 in the rearward end of the hammer 60 and bears against a thrust washer 88 at the back of the recess. The cam bushing or drive member 7S is also formed with a p air of circumferential slots 89 having suitable spring members or plates (not shown) inserted therein for holding the retaining balls 86 in place. The rearward annular end of the bushing 78 is received telescopically within the cup-shaped forward portion of the member 70 and is provided with a pair of endwise complementary V-shaped cam tracks or grooves 91 (Fig. 6) having lowermost points 92 and having raised circumferential bearing surfaces 93 extending between the adjacent ends of the grooves 91 and merging with the latter. The` V-shaped cam tracks 91 are disposed in opposed or inverted relation to the V-shaped cam tracks Sti on the member 70 with the ball elements 79 being disposed therebetween for rolling movement in the channels formed by the opposed cam tracks or grooves. Extending rearwardly from the same axial end of the bushing 78 and spaced radially inwardly from the cam tracks 91 are a pair of axial extensions 94 adapted to seat at their ends against an annular portion 96 at the rearwardmost end of the member 70 with the bearing surfaces 82 and 93 also in contact with each other. The side edges of the extensions 94 provide suitable axially extending shoulders or abutments 97 for the opposite ends of the cam tracks 91. As will hereinafter appear, the cam ball elements 79 roll in the channels formed by the opposed cam tracks 80 and 91 and the rolling movement of the balls 79 is limited by engagement of the balls at the opposite ends of the cam tracks with the cooperating stops or abutments 83 and 97.

As shown clearly in Figs. 2, 5, 6, and 9 of the drawings, the hammer 60 is provided with two diametrically opposed and aligned ears or projections 101 which extend rearwardly in parallel relationship up within the spring '71 and on each of the opposite sides of the cam bushing 78 and the cam lifter plate 70. The inner and outer surfaces of the hammer ears 101 are curved so as to conform to the curvature of the spring coils and the member 70 but they are spaced from the spring 71 to permit free relative movement between the spring and the hammer 60.

The inner surfaces of the hammer ears 101 are also in slidable engagement with the opposite outer sides of the member 70.

The hammer ears 101 thus serve to maintain the parts of the impact unit in perfect alignment and insure perfect `alignment of the slots 65 of the bifurcated portions 69 of the cam lifter plate 70 with the axial hammer openings 63. As a consequence, the impact rods 64 are always maintained in perfect and absolute alignment within the hammer openings 63 for free relative slidable axial movement to and from impacting position. Since the parallel inner surfaces of the hammer ears 101 are parallel with the axial openings 63 of the hammer 60 and bear against the opposite parallel surfaces of the cam lifter plate 70, absolute parallelism between the parts is maintained at all times. Any forces tending to rotate the hammer 60 in one direction relative to impact rods 64 so as to tend to effect their disalignment and to cause the rods to bear against one side of the hammer slots 63, also, through the hammer ears 101, tend to rotate the rod-carrying cam lifter plate 70 in the same direction and with equal force, thus preventing any such disalignment or binding engagement. The hammer ears 101 also guide the cam lifter plate 70 in a true axial direction when it is cammed upwardly by the cam bushing 78 and serve to restrain the plate 70 against rotative movement.

On the under side of the housing section 21 is a switch grip handle 102 (Fig. l) having a trigger block 103 slidably received within an opening in the front of the handle for closing and opening the electric circuit to the motor, as described in detail in the earlier application Serial No. 81,966. The trigger block 103 may be held in its inward position for continuous operation of the tool by a spring pressed locking plunger 104 extending through the forward portion of the handle and may be quickly unlocked by depressing and releasing the trigger block 103.

When a nut or bolt is to be tightened, the operator first places the socket (not shown), secured on the square 57, over a nut or bolt head and then presses the trigger block 103 inwardly to complete the electric circuit and effect operation of the tool motor. As the armature shaft 26 rotates, the pinion gear 46 on its forward end rotates the planetary gears 47, causing them to move bodily around within the ring gear 52 in a counterclockwise direction and thus causing rotation of the spider 49 and the drive shaft 50. This driving torque is further transmitted to the inpact rods 64 and the hammer 60 through the cam bushing 78 and the cam lifter plate 70.

At the outset of the tightening operation, the impact rods 64 are maintained in projected position within the slots 66 of the impact spindle 56, as shown in Fig. 1, and abut against the respective ends thereof so as to rotate the impact spindle 56 and the socket and tighten the nut or bolt. As long as the nut or bolt is relatively free running, the impact rods 64 remain in constant engagement with the impact spindle 56 and the cam balls 79 are disposed at the low points 84 and 92 of the cam tracks 80 and 91 on the elements 70 and 78, respectively. However, when the nut or bolt approaches tightened position and resistance to rotation is suicient to overcome the driving torque of the tool motor and the expansive force of the spring 71, then the unitary rotation of the impact spindle 56, impact rods 64, hammer 60, and the cam lifter plate 70 is arrested. But since the spindle 50 and the cam bushing 78 continue to rotate under the driving force of the motor, the cam balls 79 are caused to roll to the high points of the cam tracks 80 and 91 thereby shifting the cam lifter plate 70 rearwardly in a straight longitudinal or axial direction under the guiding inuence of the hammer ears 101 and against the force of the spring 71. Since the cam lifter plate 70 carries and supports the impact rods 64, these rods also are moved rearwardly relative to the hammer 60 and they are thus withdrawn from the slots 66 to a point slightly above the rearward surface of the then stationary impact spindle 56.

ln this withdrawn rearward position, the impact rods 64 are out of driving engagement with the impact spindle 56 and hence are free to rotate under the driving torque of the drive shaft 50. Consequently, the impact rods 64 and the hammer 60 resume rotation in a clockwise or tightening direction, pass over the narrow lands between the impact spindle slots 66, and then are projected under the urge of the spring 71 back into the slots 66 so that the impact rods 64 will be impacted against the far end of the slots 66 upon continued rotation. From the time that driving engagement is broken oft', as above described,- until the rods 64 impact against the ends of the slots 66, the drive shaft 50 accelerates under the driving force of the motor and the compression of the spring 71 at an ever increasing rate so that the impact rods 64 strike the impact spindle 56 with great force. Because of the solid, massive construction of the hammer 60, this impacting force is of relatively great magnitude.

Because of the fact that the impact rods 64 are circular and are slidably received in the slots 65 of the cam litter plate '70, they are free to rotate relative thereto. As a result of this rotative connection, the rods 64 are rotated slightly as they are moving into impacting position against the far ends of the slots 66, thus presenting diiTerent circumferential portions of the lower or blowdelivering ends of the rods to the slot abutments for each blow. The rods, therefore, present new wearing surfaces and have a long operable life.

It is to be noted further that when the driving engagement is initially broken ott between the impact rods 64 and the impact spindle 56, the cam balls 79 engage the cam elements 78 and 70 at the high points of theV respective V-shaped cam tracks 91 and 80. Consequently, as the impact rods 64 again accelerate toward impacting engagement with the impacting spindle 56, the heavy spring 71 will force the cam lifter plate 70 axially forwardly, the cam balls 79 moving downwardly into the low points of the cam tracks and thereby eifecting rapid rotation of the member 70. The cam lifter plate 70 is thus caused to accelerate ahead of the drive shaft 50 and the impact rods 64 are caused to strike the impactingl spindle 56 with an increased impacting force and at a rate of rotation over and above that of the drive shaft 50. After impact, which causes further limited rotative and tightening movement of the impact spindle 56, the impact rods 64 are again withdrawn as above described and are again rotated relative to the impact spindle 56 and thrown again into impacting engagement with the spindle 56 at the far ends of the next succeeding slots 66. The impact rods 64 thus strike the impact spindle 56 two impacting blows for every revolution of the hammer 60 and these blows may be repeated until the work has been tightened to the desired degree.

During the operation of the tool when the impacting blows are being delivered to the work, the impact rods 64 are freely slidable within the hammer slots 63 to and from spindle engaging position since they are maintained in perfect alignment therewith by the hammer ears 101 which engage the sides of the cam lifter plate 70. The hammer ears 101 also by their slidable engagement with the sides of the cam lifter plate 70 cause that plate to move in a truly longitudinal direction and parallel to the inner sides of the hammer ears. Since the inner Sides of the hammer ears 101 are also parallel with the hammer slots 63, the sliding movement of the rods 64 is also parallel therewith. Consequently, it is apparent that by the arrangement of the parts they are maintained in the desired alignment so that their free operation is achieved at all times and no binding of the rods 64 in the hammer slots 63 results.

When it is desired to utilize the tool for the purpose of loosening a nut or bolt, the operator need only rotate the plastic reversing ring 28 to its other position as above described, and the electric motor will thus operate in the reverse direction and the drive shaft 50 will also be rotated in the reverse direction. Since the work is in a tightened state, the impact rods 64 will immediately commence imparting blows against the spindle S6 until the resistance to rotation otfered by the work is less than the` driving torque of the cam lifter plate 70, at which point the impacting blows cease and the impact rods 64 remain in constant free running engagement with the impact spindle 56. The tool may then be allowed to remain inr 7 engagement with 4the work until `.the work is entirely disengaged :from a threaded opening.

In :Figs 2, 3, 7, and k8, 'the operating positions of the rotatable lca'rn elements 78 and '70 and the cam balls '79 corresponding to the engaged and disengaged positions of the impact clutch mechanism are shown in detail. As seen in'Pigs. 2 and 7, the balls "'79 are seated -in the lowermost points 92and 84 ofthe respective cam tracks 91 and 8l), the cup-shaped central portion of the cam lifter plate 70 thereby receiving the drive cam member or bushing 78 in retracted position .therein with the impact rods ed in constant engagement with the ends of 'the slots 66 in the anvil 56. This l.position of the cam elements constitutes the normal driving or free running condition of the .tool 'm which :there is no substantial resistance torque of the work piece tending to prevent rotation of the anvil S6. Under fsuch free running conditions with the work piece exerting only'a negligible resistance torque, the cam balls 79 provide the necessary continuous driving connection between the driving cam member or bushing 78 on the shaft 50 and .the .driven cam element 70 which is in turnoperatively connected through the impact rods 64 and the hammer 60 to the anvil 56. From the sectional view of Fig. 2, it will be seen that the lateral extent or width of the respective cam tracks 80 and 91 radially of the members 70 and '78 is su'icient to provide a high degree of circumferential surface contact with the balls 79 so that the latter are not subjected to undesirable shearing stress. More specifically, it will be seen that the areas of surface contact kof the cam tracks with the cam balls 79 extend radially on each side of the center lines or axes of the cam balls so that the longitudinally acting forces exerted on the cam balls 79 are primarily compressive in contradistinctionto the undesirable shear conditions which would prevail if the areas of surface Contact of the cam tracks with the cam balls were confined to oppositely disposed and axially ofset portions of the cam balls.

In Figs. 3 and 8 the cam elements are shown in position for effecting declutching of the impact clutch mechanism in response to a high resistance torque of the work piece 'tending to prevent rotation of the anvil 56 and the interconnected cam lifter plate 7i). Thus, it will be seen that the cam bushing 78 has continued to rotate relative to the cam element 70 with the result that the cam balls 7-9 roll upwardly along the inclined sides of the cam tracks :91 and 89, the extent of rolling movement of the balls .79 .being limited by the axial abutments or shoulders 83 and 97 at the ends of the cam tracks 80 and 91, respectively. This movement of the cam elements causes the cam lifter plate 70 to be shifted axially and rearwardly relative to the shaft t) so that the cam bushing 73 is in slightly extended position relative to the cup-shaped portion of the cam element 70 (Fig. 3). lt is a highly important .feature of my invention that even in this latter position of the camelements, the area of surface contact or bearing area of the cam track 93 with the cam balls 79 still extends klaterally or radially on opposite sides of the center lines of the balls 79. Similarly, the area of surface contact-of the cam track Sti with the balls 79 also extends on `opposite sides of the center lines of the cam balls whereby the balls 79 are still subjected primarily to compression rather than shear even in the extreme positions of the cam elements.

As'hereinbefore mentioned, the avoidance of shear conditions is `particularly important in an impact clutch mechanism of the present character since it will be readily understood that the cam balls 79 and the grooves Stb and L91 must withstand the forces developed upon repeated impact blows of the clutch mechanism. It has been found that when ball cam elements and coacting cam tracks or grooves are so arranged that the balls are subjected to shear there is marked tendency for the corner edges of the grooves to break oit as a result of the concentration of forces atspaced points of the bearing areas. However, the cam'tracks and the-cam balls in the-*present invention are so arranged that the effective contact -or bear-ing areas between the tracks and the balls at 'the opposite sides of the "latter and radially thereof always overlap radially to a substantial extent Athereby providing adequate distribution of the impact forces and avoiding a condition in which oppositely directed-forces are'exerted on `the -balls in parallel spaced planes which would result in placing the balls under excessive shearing stress. Also, as best seen in Figs. 2 and 3, the surfaces of the cam tracks or channels and 91 have a transverse curvature conforming to the curvature-of the balls 79 radially thereof so that there is always line contact between the balls and the-surfaces of the cam tracks. As a result the balls are fully guided during rolling movement along the cam tracks and the stress or load on the tracks is effectively distributed rather than being concentrated at only .one or a few points.

An additional advantage of the cam arrangement of the Ypresent invention resides in the generally telescopic arrangement of the cam elements whereby the cam bushing 73 is received within and-substantially enclosed by the skirt portion 81 (Figs. 2 and 3) of the cam member 79. The cam balls 79 are thereby enclosed to an appreciable degree at all times by the arrangement of the hammer 60 and the coacting cam members 78 and 70 with the beneficial result that grease or other lubricant is readily retained in the operating or bearing areas of the cam -balls and cam tracks. In addition, the enclosed condition of the cam balls and tracks serves to protect the latter yfrom wear particles-of metal and other foreign particles which would interfere with proper operation of the cam mechanism.

It is to be further understood that while the preferred embodiment of the present invention has been described above as an impact wrench, the very same tool by means of special attachments secured on the square 57 may be adapted for other uses such as a screw driver, tapper, reamer, drill, sander, polisher or wire brush driver. When utilized -for ythese latter purposes, the tool will be primarily free running without operation as an impacting wrench. However, such impacting features are permitted to become effective when used as a screw driver or when used as a drill and the drills or bits become frictionally held against rotationbefore the drill hole has been cornpleted. The impacting blows of the hammer rods 64 then are applied to the drills and bits to overcome such frictional resistance and any overload on the motor is eliminated.

Although there has been described above and illustrated in the drawings a preferred embodiment of an impact tool comprising the present invention, it is apparent that modifications and changes may be made in the details of structure and mode of operation without departing from the spirit and scope of the lappended claims.

I claim:

11. In a .rotary impact tool of the type having rotatable hammer and anvil members and impact means carried by the hammer member and normally engaging the anvil member but being disengageable therefrom, cam means comprising a rotary driving member having inclined cam tracks, a rotary driven member operatively connectible with the hammer and impact means and having complementary inclined cam tracks disposed in coacting relation with the first-named cam tracks, said cam tracks extending in arcuate paths circumferentially of the respective driving and driven members and providing therebetween a pair of opposed parallel cam surfaces, and a plurality of cam balls rollably disposed between said cam surfaces for unconned rolling movement along the arcuate length of said cam tracks, said cam surfaces having transverse curvature conforming to the curvature of the balls along the entire arcuate length of the tracks whereby said balls are always in line contact with said cam surfaces for guiding said balls vduring said rolling movement and for effective distributienet `stress-on the cam tracks, and the contact 9 areas of said cam surfaces with said balls extending in mutually overlapping relation radially of the balls whereby the load on the balls is solely in compression and not in shear thereby avoiding detrimental edge loading of the cam tracks.

2. In a rotary impact tool having coacting rotatable hammer and anvil members, impact means carried by the hammer member for striking the anvil member, a spring normally urging said impact means into engagement with the anvil member, and driving means for the hammer member; the improvement comprising cam means for connecting said driving means with the hammer member and for disengaging said impact means from the anvil member, said cam means comprising a pair of driving and driven cam members operatively connectible to said driving means and to said hammer respectively, said cam members having a plurality of complementary inclined cam tracks extending in arcuate paths circumferentially of said cam members and providing therebetween opposed parallel cam surfaces, and a plurality of cam balls rollably disposed between said cam surfaces for unconfned rolling movement along the arcuate length of said cam tracks, said cam surfaces having transverse curvature conforming to the curvature of the balls along the entire arcuate length of the tracks whereby said balls are always in line contact with said cam surfaces for guiding said balls during said rolling movement and for effective distribution of stress on the cam tracks, and the contact areas of said cam surfaces with said balls extending in mutually overlapping relation radially of the balls whereby the load on the balls is solely in compression and not in shear thereby avoiding detrimental edge loading of the cam tracks.

3. In a rotary impact tool of the type having rotatable hammer and anvil members and impact means carried by the hammer member and normally engaging the anvil member but being disengageable therefrom, cam means comprising a cup-shaped driven member operatively connectible to the hammer and impact means, a driving bushing extending axially into said driven member, the axial inner end of said bushing having endwise inclined cam tracks and the base of said cup-shaped member having complementary inclined cam tracks at its inner face disposed in coacting relation with the cam tracks on said bushing, said cam tracks extending in arcuate paths circumferentially of said bushing and said driven member and providing therebetween a pair of opposed parallel cam surfaces, and a plurality of cam balls rollably disposed between said cam surfaces for unconined rolling movement along the arcuate length of said cam tracks, said cam surfaces having transverse curvature conforming to the curvature of the balls along the entire arcuate length of the tracks whereby said balls are always in line contact with said cam surfaces for guiding said balls during said rolling movement and for effective distribution of stress on the cam tracks, and the width of said cam tracks radially of said bushing and of said driven member being suicient to provide mutually overlapping areas of contact between said cam surfaces and said balls radially of the latter whereby the load on the balls is solely in compression and not in shear thereby avoiding detrimental edge loading of the cam tracks.

4. In a rotary impact tool of the type having rotatable hammer and anvil members and impact means carried by the hammer member and normally engaging the anvil member but being disengageable therefrom, cam means comprising a rotary driving bushing having inclined cam tracks formed in one axial end thereof, a rotary cup-shaped driven member operatively connectible with the hammer and impact means and having a laterally directed wall at one end thereof and an axially extending skirt portion adapted to receive therein said one axial end of said bushing, said laterally directed wall having complementary inclined cam tracks at its inner face in coacting relation with said first-named cam tracks, said cam tracks extending in arcuate paths circumferentially of said bushing and said driven member and providing therebetween a pair of opposed parallel cam surfaces, and a plurality of cam balls rollably disposed bet. een said cam surfaces for uncontned rolling movement along the arcuate length of said cam tracks for drivingly connecting said bushing .and said driven member when said balls are disposed at the lowermost points of said cam tracks and for shifting said driven member axially to disengage the impact means from the anvil member when said balls are at the uppermost points of said cam tracks, said cam surfaces having transverse curvature conforming to the curvature of the balls along the entire arcuate length of the tracks whereby said balls are always in line contact with said cam surfaces for guiding said balls during said rolling movement and for effective distribution of stress on the cam tracks, and the width of said cam tracks radially of said bushing and of said driven member being sucient to provide mutually overlapping areas of contactbetween said cam surfaces and said balls radially of the latter whereby the load ou the balls is solely in compression and not in shear thereby avoiding detrimental edge loading of the cam tracks.

5. In a rotary impact tool of the type having rotatable hammer and anvil members and impact means carried by the hammer member and normally engaging the anvil member but being disengageable therefrom, cam means comprising a rotary driving bushing having V-shaped cam tracks formed in one axial end thereof, a rotary cup-shaped driven member operatively connectible with the hammer and impact means and having a laterally directed wall at one end thereof and an axially extending skirt portion adapted to receive therein said one axial end of said bushing, said laterally directed wall having complementary V-shaped cam tracks at its inner face in inverted relation to said first-named cam tracks, said cam tracks extending in arcuate paths circumferentially of said bushing and said driven member and providing therebetween a pair of opposed parallel cam surfaces, and a plurality of cam balls rollably disposed between said cam surfaces for unconined rolling movement along the arcuate length of said cam tracks for drivingly connecting said bushing and said driven member when said balls are disposed at the lowermost points of said cam tracks and for shifting said driven member axially to disengage the impact means from the anvil member when said balls are at the uppermost points of said cam tracks, said cam surfaces having transverse curvature conforming to the curvature of the balls along the entire arcuate length of the tracks whereby said balls are always in line contact with said cam surfaces for guiding said balls during said rolling movement and for effective distribution of stress on the cam tracks, and the Width of said cam tracks radially of said bushing and of said driven member being sufficient to provide mutually overlapping areas of contact between said cam surfaces and said balls radially of the latter whereby the load on the balls is solely in compression and not in shear thereby avoiding detrimental edge loading of the cam tracks.

6. The structure of claim 5 further characterized in that said skirt portion of said driven member is formed with axially extending shoulder portions disposed at the ends of the cam tracks in said driven member for providing abutments for said balls, and said bushing is formed with axial extensions at Said one end projecting axially beyond the cam tracks in said bushing for providing additional cooperating abutments for said balls.

References Cited in the file of this patent UNITED STATES PATENTS 1,585,140 Erban May 18, 1926 1,634,861 Weymann July 5, 1927 1,657,274 Niedhammer Jan. 24, 1928 1,881,633 Johnson Oct. 11, 1932 2,585,486 Mitchell Feb. 12, 1952 2,587,712 Dodge Mar. 4, 1952 

